BX RTJ 2

[nctexan22]

Because BX gaskets are pressure energized seals, does this qualify them as "self-energized" as defined by Section VIII Div. 2? Table 4.16.1 (gasket factors for determining bolt loads) gives "m" and "y" as 0 for self-energizing gaskets, but the seating stress for ring joints is very high. Which factors apply for a BX RTJ gasket?

It seems to me that if it's a self-energized seal, the internal system pressure would increase the gasket seating stress on the flange contact surface, though Section VIII tells us that the axial force required to compress this type of gasket is zero. I'm a bit confused.

 

[gasketguru]

Whilst BX are thought of as “pressure energized”, they are pretty much interference—fit energised initially as they start slightly oversize compared to the groove and coin-in under compression.

These joints were originally designed in the 1950’s using a modified Taylor-Forge method. They were assumed to be pressure energised, with the belief that to get a good seal the stress on the outer flanks should be 3x the system pressure – hence the bevelled corners are 1/6 of the height of the ring gasket. The flanges were designed around 75,000 psi yield material and a 52,500 psi bolt stress.

In terms of the load factors the traditional Taylor-Forge components of H (hydrostatic force) and Hp (gasket reaction) were H=pi/4 x G^2 x P where G is the seated (compressed) O/D of the ring itself and Hp is pi x G x Lg x P tan phi, where Lg is the height of the ring, P is the pressure and phi is the angle of the groove i.e. 23 degrees. The original design rules also allowed for the raised face diameter outboard of the groove to carry the total bolt load without exceeding 40% yield on the flanges.

Hope this helps.

 

[TWP]

Basics design is outlined in ASME Paper No. 63-PET-3 June 4, 1963 "design of high pressure integral and welding neck flanges with pressure energized ring joint gaskets (R. Eichenberg). This doc. can be obtained from Linda Hall lib (

http://www.lindahall.org).

By the way...did you finish your design template for this particular type of BX gaskets?
I'm struggling with some details on loads Wm1, Wm2 and factors y,m (Wm1,Wmw,y,m as per ASME) for BX gaskets and would welcome some support.

 

[gasketguru]

I’m not entirely sure I fully understand your question here – I have a copy of the 63-PET-3 paper, which followed on from the 57-PET-23 one ( I couldn’t find them on the link you mentioned though)

Basically, there is no m and y as such, with Wop being H+Hp of course and H is the usual pi/4*G^2P and Hp is pi*G*Lg*Ptan phi where Lg is the ring height and the angle phi is the 23 degree groove angle. – Again the ring bevel is assumed as 1/6 of the ring height. This paper also introduced the concept of under-square section seals based on their relationship of the height to width relative to the pressure vs. the 15ksi standard ones. Otherwise the rest was conventional Taylor-Forge protocol. If you have the paper, then this detail is all explained for you of course. - Let me know the bit you are missing and I'll try to assist...

 

[TWP]

Thanks Gasketguru. My company got an official copy of Paper 63-PET-3 from Linda Hall lib. That's what I wanted you to know. Unfortunately we could'nt get a copy of Paper 57-PET-23. Perhaps you know where to obtain a copy of this paper???

On basis of Paper 63-PET-3 and the conventional Taylor Forge model (as per "Modern flange design bulletin 502")I did prepare a set of bolt+flange (Excel) design rules.

This was actually not a major task.
The only thing I'm missing in the theory is bolt load Wm2.
This bolt load is 'normally'always required in most design codes.
Perhaps we don't need it here because it's too small compared to Wop (Wm1)??
I know that some professionals in the USA use y-values as listed in the ASME tables.
If you have any relevant info on the derivation of a logic formula for Wm2 and therein used y-value and gasket width b, then please inform me.
Thanks in advance. TWP (The Netherlands)

 

[gasketguru]

If you look at the paper closely, you will see no mention of Wm1 or Wm2 as such, because no “initial seating load” is considered due to the fact that these are thought of as “pressure energised” gaskets (or as I say: “interference-fit energised”…)

Thus, there is an initial gap of 1/8” between the flanges which is closed as it is bolted-up, and no specific detail is mentioned about what percentage of the overall bolt load might be taken to do this – I assume the high hydrostatic force is assumed to be far greater of course. Similarly “b” is not relevant as you are assuming the load reaction on the outer flank of the seal (seated ring diameter etc. in the equations).

The 1957 paper was the design basis for the 150 series rings which are all square in section and the later one for many of the 160 series, though there are still a few which do not obey either of these rule sets so maybe I am missing a later AWHEM paper for some of the others.